Some engines disposed in a vehicle are of a type called a cylinder direct injection engine. This type of an engine has a spark plug provided in a cylinder head at a substantially central portion of a combustion chamber. The combustion chamber is formed between the underside of the cylinder head and a top surface of a piston. In addition, intake and exhaust valves are disposed in the cylinder head. The intake valve is positioned on one side of the cylinder head, but the exhaust valve is located on the other side thereof. Further, an injector is disposed in the cylinder head on one side thereof. The injector expels fuel in the form of a conical stream onto the top surface of the piston.
One such example of a cylinder direct injection engine is disclosed in published Japanese Patent Application Laid-Open No. 7-217478. A fuel injection controller for use in a spark ignition engine of a cylinder injection type as disclosed in this publication controls knocking without detracting from fuel efficiency when the knocking occurs.
Another example is disclosed in published Japanese Patent Application Laid-Open No. 9-144543. A direct injection spark ignition internal combustion engine as disclosed in this publication has a spark plug disposed at a substantially central portion of a combustion chamber. In addition, an intake port is positioned offset from the axis of a cylinder. Further, a fuel injection valve is provided at a depression angle (.theta.=30.degree..+-.10.degree.) at the intake port adjacent to an opening of the combustion chamber in such a manner that cone angle .alpha. of fuel spray is set to be 70.degree..+-.20.degree.. This construction provides improved engine performance.
A further example is disclosed in published Japanese Patent Application Laid-Open No. 10-288127. In a combustion chamber of an internal combustion engine as disclosed in this publication, a concave portion is provided on a cylinder head wall surface on the side of a spark plug near a nozzle aperture of a fuel injection valve. When the fuel injection valve discharges fuel therefrom under atmospheric pressure conditions outside the combustion chamber, then a first angle .theta.1 is defined between a central axis defined by the fuel spray and a peripheral surface of the fuel spray. In this state, a second angle .theta. defined between a slanted wall surface at the concave portion and the conical surface of the fuel spray is set to be greater than the first angle .theta.1. The conical surface of the fuel spray defines first angle .theta.1 with respect to the central axis of the fuel spray from the fuel injection valve after the same is fitted to the combustion chamber. Such a construction restrains or minimizes the occurrence of an attractive force, which otherwise would attract atomized fuel to the cylinder head wall surface. In addition, fuel is combusted in a stable manner, which results in a reduced amount of soot.
A further example is disclosed in published Japanese Patent Application Laid-Open No. 10-339145. A cylinder direct injection type of a spark ignition engine as disclosed in this publication includes a swirl control valve for imparting swirl to intake air that enters a cylinder through an intake port. Further, a pattern of atomized fuel driven out of an injector is set to have a hollow, conical shape that is in a state of initial atomization. In addition, a cap portion of a piston is formed with a cavity that is dented so as to receive the atomized fuel. This structure provides a combustion chamber adapted for this type of an engine.
A cylinder direct injection engine heretofore employed is designed to inject fuel directly into a combustion chamber from an injector during a compression stroke, thereby forming a stratified air-fuel mixture. For this reason, a large number of improvements in a piston pattern have been made.
However, injection timing of the injector as well as a piston phase influence the formation of the stratified mixture. In addition, the injection timing is limited to a narrow range.
FIG. 26 illustrates a cylinder direct injection engine (hereinafter simply called an "engine") 102. The engine 102 has a 75 mm cylinder bore and a 90 mm piston stroke. FIG. 27 is a cross-sectional view, showing a spark plug portion of the engine 102 in a 30.degree. crank angle (CA) BTDC (Before Top Dead Center) state. FIG. 28 is a cross-sectional view, illustrating the spark plug portion of the engine 102 in a 45.degree. crank angle (CA) BTDC state. FIG. 29 is a cross-sectional view, illustrating the spark plug portion of the engine 102 in a 60.degree. crank angle (CA) BTDC state.
The engine 102 has a piston 108 disposed in a cylinder block 104 for reciprocation therein. In addition, a concave area 132 is defined at a top surface 108a of the piston 108.
The piston 108 is connected to a crankshaft (not shown) through a connecting rod (not shown). A combustion chamber 116 is formed between the underside of a cylinder head 106 and the top surface 108a. The cylinder head 106 is positioned on an upper surface of the cylinder block 104. A spark plug 118 is provided in the cylinder head 106 at a substantially central portion of the combustion chamber 116. An injector 120 is located in the cylinder head 106 on the intake side of the cylinder head 106. The injector 120 permits fuel to be expelled therefrom in the form of a conical stream directly into the combustion chamber 116.
The injector 120 injects the fuel therefrom into the concave area 132 during a second half of the compression stroke. At this time, atomized fuel "F" (FIG. 27) is captured at the concave area 132, and is then collected near the spark plug 118. Such collected fuel "F" forms a stratified mixture in cooperation with a lean mixture that surrounds the stratified mixture.
A certain period of time for the fuel to be atomized must be maintained between fuel injection and ignition. Such a period of time has requirements that vary, depending upon engine speed and engine load (i.e. a longer period of time between injection and ignition elapses for lower loads and speeds, and a shorter period of time between injection and ignition elapses for higher loads and speeds). Typically, ignition timing covers a range between a 30.degree. and a 60.degree. crank angle (CA) BTDC.
In the disclosed engine 102, however, the 60.degree. crank angle (CA) BTDC as illustrated in FIG. 29 results in an improper positional relationship between the concave area 132 and the atomized fuel "F". This causes inconveniences of insufficient fuel capture, unsatisfactory stratification, and unstable combustion, which are all disadvantageous in view of practical use.
FIG. 30 discloses an engine 202 having a reduced-diameter cylinder bore. In the engine 202, a greater portion of atomized fuel "F" is shown adhered to a cylinder wall inside a cylinder block 204. This undesirable condition leads to scuffing which is caused by oil film deposited on a wall surface of the cylinder block 204, which is disadvantageous in view of practical use.
FIG. 31 illustrates another cylinder direct injection engine 302 heretofore employed. In this engine 302, fuel is injected from an injector 320 during a compression stroke, and is then delivered to a location near a spark plug 318 by means of a tumble flow (or swirl flow), thereby forming a stratified mixture.
At this time, when a piston stroke is small or short, then the tumble flow (or swirl flow) is reduced in strength, with a concomitant deficiency in fuel delivery. This brings about yet further inconveniences of an unsuccessful stratified mixture and thus unstable combustion.
Moreover, when the cylinder bore is too large (in diameter) in each of the above-described engines, except for the engine having the reduced cylinder bore of FIG. 30, a still further convenience is encountered. More specifically, as illustrated in FIG. 32, a richer mixture gathered near a spark plug 418 of an engine 402 and a leaner mixture surrounding the richer mixture are distributed incompletely, which results in unstable combustion.
In the above engine 402, when a stroke-bore ratio is extremely great or small (see FIG. 33), then there occurs yet another inconvenience in that flame is spread in a non-uniform fashion during combustion, with consequential unstable combustion.
FIG. 34 shows a cylinder direct injection engine 502 having a 30.degree. cone angle, a 82 mm cylinder bore of a cylinder block 504, and a 45.degree. injector installation angle of an injector 520.
In the engine 502 constructed according to the above settings, atomized fuel "F" jetting from the injector 520 during an intake stroke is spread in a combustion chamber due to various factors such as energy during fuel injection, fuel reflection on a top surface of a piston, fuel reflection on a cylinder wall, a flow of air, and fuel vaporization due to receipt of heat. However, since the cylinder bore diameter is excessively large with respect to the cone angle, then atomized fuel "F" is insufficiently spread in the combustion chamber, even when the spark plug is ignited. Such insufficient fuel dispersion causes still further inconveniences in that an air-fuel mixture is distributed in a non-uniform manner, with ensuing unstable combustion.
Just for reference, various configurations or patterns of the concave area 132 formed at the top surface 108a of the piston 108 are illustrated in FIGS. 35-38.
A piston 108-1 as shown in FIGS. 35(a)-35(c) has a rectangular-shaped concave area 132-1 formed on a top surface 108a-1 thereof. The concave area 132-1 extends from the intake side to the exhaust side.
A piston 108-2 as shown in FIGS. 36(a)-36(c) has a rectangular-shaped concave area 132-2 formed on a top surface 108a-2 thereof. The concave area 132-2 extends from the intake side to the exhaust side. In addition, protruding wall portions W1, W2 are formed at both ends of the concave area 132-2 in a direction between the intake side and the exhaust side. The protruding wall portions W1, W2 project in an upward direction from the piston 108-2, and thereby define a depth of the concave area 132-2.
A piston 108-3 as shown in FIGS. 37(a)-37(c) has a rectangular-shaped first concave area 132-3 defined on a top surface 108a-3 thereof. The first concave area 132-3 extends from the intake side to the exhaust side. In addition, protruding wall portions W1, W2 are formed at both ends of the first concave area 132-3 in a direction between the intake side and the exhaust side. The protruding wall portions W1, W2 project upwardly from the piston 108-3, and thereby define the depth of the first concave area 132-3. Further, a rectangular-shaped second concave area 133-3 is formed at a central portion of the first concave area 132-3. The second concave area 133-3 extends from the intake side to the exhaust side.
A piston 108-4 as shown in FIGS. 38(a)-38(c) has a rectangular-shaped first concave area 132-4 defined on a top surface 108a-4 thereof. The first concave area 132-4 extends from the intake side to the exhaust side. In addition, protruding wall portions W1, W2 are formed at both ends of the first concave area 132-4 in a direction between the intake side and the exhaust side. The protruding wall portions W1, W2 project upwardly from the piston 108-4, and thereby define a depth of the first concave area 132-4. Further, a rectangular-shaped second concave area 133-4 is formed on the first concave area 132-4 adjacent the exhaust side thereof. The second concave area 133-4 extends from the intake side to the exhaust side.
In order to obviate or at least minimize the above-described inconveniences, one aspect of the present invention provides a cylinder direct injection engine including a spark plug positioned in a cylinder head at a substantially central portion of a combustion chamber, the combustion chamber being formed between the underside of the cylinder head and a top surface of a piston, intake and exhaust valves disposed in the cylinder head on opposite sides thereof, and an injector provided in the cylinder head on one side of the cylinder head for injecting fuel in the form of a conical stream onto the top surface of the piston, wherein a difference in size between a piston stroke and a cylinder bore diameter ranges from about 0% to about 4%, and the piston stroke is greater than the cylinder bore when the difference is unequal to 0%.
Another aspect of the present invention provides a cylinder direct injection engine including a spark plug positioned in a cylinder head at a substantially central portion of a combustion chamber, the combustion chamber being formed between the underside of the cylinder head and a top surface of a piston, intake and exhaust valves disposed in the cylinder head on opposite sides thereof and an injector provided in the cylinder head on one side of the cylinder head for injecting fuel in the form of a conical stream onto the top surface of the piston, the improvement wherein the engine has a piston stroke of about 50 mm to about 80 mm, cylinder bore diameter of about 50 mm to about 92 mm, and a stroke-bore ratio of about 0.8 to about 1.4.
A further aspect of the present invention provides a cylinder direct injection engine including a spark plug positioned in a cylinder head at a substantially central portion of a combustion chamber, the combustion chamber being formed between the underside of the cylinder head and a top surface of a piston, intake and exhaust valves disposed in the cylinder head on opposite sides thereof and an injector provided in the cylinder head on one side of the cylinder head for injecting fuel in the form of a conical stream onto the top surface of the piston, wherein a cylinder bore diameter is set to be less than about 85 mm with respect to a cone angle of about 15.degree. to about 90.degree. of fuel injected in a conical stream pattern from the injector.
As discussed above, the cylinder direct injection engine according to one aspect of the present invention is constructed in such a manner that a difference in size between the piston stroke and the cylinder bore diameter ranges from about 0% to about 4%, and the piston stroke is greater than the cylinder bore when the difference is unequal to 0%. This construction provides a reduced amount of displacement of the piston with respect to a crank rotational angle, and thus an ideal stratified mixture during compression stroke injection in a wider range. As a result, fuel is combusted in a stable manner.
In addition, the cylinder direct injection engine according to another aspect of the present invention is configured to have a 50 to 80 mm piston stroke, a 50 to 92 mm cylinder bore, and a 0.8 to 1.4 stroke-bore ratio. Such a construction is possible to prevent scuffing, and thus to provide stable combustion.
Further, the cylinder direct injection engine according to a further aspect of the invention is configured to have the cylinder bore set to be less than 85 mm with respect to the 15.degree. to 90.degree. cone angle of fuel expelled in the form of a conical stream from the injector. Such a construction stabilizes combustion.